Hydraulic wheel motor and pump

ABSTRACT

A hydraulic wheel motor and pump are provided for use with a motor vehicle having a pair of mechanically driven wheels and a pair of non-mechanically driven wheels. The hydraulic pumps cooperate with the mechanically driven wheels, and the hydraulic motors are coupled to the non-mechanically driven wheels. The pumps and motors are diagonally connected to one another to create a four wheel drive vehicle.

FIELD OF INVENTION

This invention relates to four wheel drive vehicles, and moreparticularly to vehicles utilizing hydraulic drive motors to drive thenonmechanically driven wheels.

BACKGROUND OF INVENTION

Hydraulic drive motors have been used on various automotive andnon-automotive vehicles for years. Hydraulic wheel motors have achieveda fair degree of commercial success in farm and off road vehicles, asshown in U.S. Pat. No. 3,584,698, assigned to Caterpillar TractorCompany. A number of hydraulic wheel motors are shown in the prior art,such as those disclosed in U.S. Pat. Nos. 2,418,123, 3,179,016,3,280,934, 3,584,698 and 3,612,205. These hydraulic wheel motors areprovided with a rotary member and a stationary member having a pluralityof cavities for pumping hydraulic fluid which periodically varies indisplacement as the stationary and the rotary member move relative toone another. Hydraulic fluid from an engine driven pump is supplied tothe hydraulic wheel motors for propelling the vehicle.

Finney, in U.S. Pat. No. 3,747,722 first suggested the use of pumpscooperating with the mechanically driven wheels of a vehicle tohydraulically drive hydraulic wheel motors coupled to thenon-mechanically driven wheels. The pumps and motors are attached to thevehicle outboard of the axle and brake drum, thereby increasing thevehicle's track by two times the pump thickness. The Finney systemcoupled front and rear wheels on the same side of the vehicle togetherhydraulically.

An object of the present invention is to provide a hydraulic driveapparatus to convert a two wheel drive to a four wheel drive vehiclewith minimal alteration to the vehicle, and without increasing thevehicle track.

Another object of the present invention is to provide a hydraulic fourwheel drive mechanism which can be utilized at low speeds and disengagedat high speeds to eliminate friction loss.

Yet another object of the invention is to provide independentlycontrollable hydraulic wheel motors or pumps adjacent each wheel whichcan be regulated to maximize tractive force in steering maneuvers.

These and other objects, feature and advantages of the present inventionwill be more fully understood with reference to the drawings,specification, and claims.

SUMMARY OF INVENTION

Accordingly, a hydraulic drive apparatus of the present invention isintended for use in a motor vehicle having a pair of mechanically drivenwheels and a pair of non-mechanically driven wheels, each of which isprovided a disc brake and caliper mechanism. The hydraulic driveapparatus includes a pair of rotary hydraulic pumps cooperating with themechanically driven wheels, and a pair of rotary hydraulic motorscooperating with the nonmechanically driven wheels. Each of the pumpsand motors are provided with a stationary member affixed relative to thebrake caliper and a rotary member affixed relative to the brake disc.The stationary member is oriented adjacent to and axially spaced fromthe brake disc forming a semi-circular segment circumaxially alignedwith the brake caliper. Fluid conduits connect the pair of hydraulicpumps to a pair of hydraulic motors causing all of the wheels to bedriven.

A variable displacement mode hydraulic motor is also described whichautomatically increases in displacement, resulting in higher activeforce upon the occurrence of a predetermined amount of driven wheelslippage.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a side elevation of a motor vehicle showing a location on thehydraulic drive apparatus;

FIG. 2 is a schematic top view of the vehicle in FIG. 1;

FIG. 3 is an enlarged partial sectional view taken along line 3--3 ofFIG. 1;

FIG. 4 is an enlarged cut-away side elevation of a hydraulic motor;

FIG. 5 is a side elevation of a hydraulic pump;

FIG. 6 is a side elevation of a hydraulic pump channel plate;

FIG. 7 is a cross-sectional view taken along line 7--7 of FIG. 6;

FIG. 8 is a side elevation of a hydraulic motor valve body;

FIG. 9 is a sectional view taken along line 9--9 of FIG. 8;

FIG. 10 is a partial schematic view of a hydraulic drive apparatusshowing the hydraulically connected pump and motor and the actuatorsystem in schematic form;

FIG. 11 is a side view of a piston roller assembly;

FIG. 12 is a right end view of FIG. 11;

FIG. 13 is a side view of timing valve assembly;

FIG. 14 is a right end view of FIG. 13;

FIG. 15 is a plot of piston velocity and position as a relationship tovalve position;

FIG. 16 is a composite plot of 4 pistons flow output pulses;

FIG. 17 is a plot of a non-mechanically driven wheel tractive forceversus fluid pressure;

FIG. 18 is a plot of tractive force versus friction coefficient for anumber of different drive conditions;

FIG. 19 is an enlarged cut-away view of the select low valve;

FIG. 20 is a load distribution diagram;

FIG. 21 is a plan view of a vehicle having an alternative embodiment ofthe invention equipped with a torque control system;

FIG. 22 is a schematic diagram for the alternative embodiment shown inFIG. 21; and

FIG. 23 is an enlarged cutaway drawing of the threshold valve.

BRIEF DESCRIPTION OF PREFERRED EMBODIMENT

FIGS. 1 and 2 show a representative motor vehicle 30 equipped with ahydraulic drive apparatus of the present invention. The motor vehicle 30is provided with a pair of front wheels 32 and 34, and a pair of rearwheels 36 and 38. The vehicle shown in the drawings has the front wheelsmechanically driven in conventional front wheel drive manner. Theinvention can also be used on rear wheel drive vehicles, and is equallyadaptable to vehicles with automatic or manual transmissions. Themechanically driven wheels, however, drive a pair of rotary hydraulicpumps mounted adjacent the conventional disc and caliper brakemechanism. Front wheels 32 and 34 are provided with rotary hydraulicpumps 40 and 42, respectively. The hydraulic pumps 40 and 42 are eachprovided with a stationary member 44 and a rotary member 46. Thestationary member is a semi-circular segment aligned with the wheel axisas shown.

The orientation of the stationary member 44 and the rotary member 46relative to the vehicle wheel assembly and disc and caliper brakemechanism can be more clearly seen in FIG. 3. Stationary member 44 isaffixed to brake backing plate 48 which is fixed relative to the spindleassembly, and also serves to support the brake caliper. Stationarymember 44 is made up of two main components; channel plate 50; andcylinder block 52. A piston assembly 54 is shown in a cylindrical bore56 and cylinder block 52. The cylinder bore is enclosed by end cap 58which defines a variable on a displacement volume bounded by thecylinder bore, the end cap and the piston assembly. The piston assemblyis provided with a roller follower which rides upon the outer peripheryof the rotary member 46. The piston assembly is also provided with analignment roller 62 which translates within a notch 64 and channel plate50. Alignment of roller 62 and notch 64 keep the piston assemblyoriented within the bore so that the roller follower axis issubstantially parallel to axis 66 about which the wheel brake rotor 68and the rotary member 46 turn.

As can be seen in FIG. 4, rotary member 46 is generally cylindrical inshape having a plurality of cam lobes 70 oriented on its outerperiphery. The rotary member 46 has an outer diameter which is limitedsufficiently so that the cam lobes will not engage the brake caliper 72,shown in phantom outline. When the vehicle is driving in a forwarddirection, rotary member 46 turns in the direction of the arrow shown.As the rotary member turns, piston assemblies 54 reciprocate withintheir respective bores. Note, the axis of the cylinder bore 74 isinclined in the forward rotation direction relative to a radial lineextending from wheel axis 66 through the center of roller follower 60 byan angle θ. θ is preferably within the range of 5° to 20°, and isselected to minimize piston friction forces, and thus reduce wear andimprove operating efficiency. Since the vehicle is run in a forwarddirection for a very high proportion of the time, inclining the cylinderbore forward in the direction of rotation, although resulting in muchhigher side load in the reverse direction, substantially reduces sideloading in the forward direction to result in a net overall wearreduction.

As the rotary member 46 rotates, piston assembly 54 reciprocate causingthe volume bounded by the cylinder bore, end cap and pistons to vary indisplacement. In order to effectively utilize this volume change to pumpfluid, an exit port is provided which communicates with the enclosedregion (shown in FIG. 10), which is connected to the timing valveassembly 78. The timing valve 78 is provided with a spool 80 andfollower 82 which translate along a valve axis 84 oriented relative toradial line 76', similar to the orientation of the piston assemblyrelative to radial line 76.

In the preferred embodiment shown, the rotary hydraulic pump 40 isprovided with eight piston assemblies mounted within the stationarymember 44 for reciprocal movement relative thereto in response to therotation of rotary member 46, which is provided with seven lobes 70.Four timing valve assemblies 78 are provided, one timing valvecooperating with two piston assemblies.

As will be described in more detail, subsequently, when the pistonassemblies move radially outwardly, the piston port is coupled to thepump output by the valve member. When the piston is moving radiallyinwardly, the piston port is coupled to the pump input by the valvemember. During a single wheel rotation, each of the eight pistons willreciprocate seven times, resulting in a total of 56 output pulses and 56input pulses. The shape and timing of these pulses is carefullycontrolled by the profile of lobes 70.

In the preferred embodiment of the hydraulic pump shown, it isspecifically designed to compactly fit within traditionally unused spacesurrounding the brake rotor and brake caliper. Typically, the brakecaliper in a disc brake system occupies a very small segment of thedisc, typically, in the order of 90°. Since the caliper is much widerthan the disc, the vehicle wheel must be designed to clear the caliperwhen it rotates. A large semi-circular toriodal segment is typicallyunutilized. As shown in FIG. 3, the rotary member 46 is affixed to thebrake disc 68 projecting axially outward therefrom. The stationarymember which is generally semi-circular shaped is mounted radiallyoutwardly from the rotary member surface adjacent to, and axially spacedfrom the brake disc 68. The rotary hydraulic pump can thereby befabricated into a small compact unit which requires little or nomodification to the vehicle wheel and the brake disc, and only slightmodification to the caliper support plate to facilitate the attachmentof the stationary and rotary members.

In operation, rotary hydraulic pumps 40 are oriented at each of themechanically driven wheels. Rotary hydraulic motors 86 and 88 cooperatewith each of the non-mechanically driven wheels. The high pressure fluidinputs to motors 86 and 88 are connected by restricted passageway 183.The limited flow is sufficient to enable the vehicle to make tight turnswithout scrubbing the tires.

The hydraulic motors are basically similar to the hydraulic pump. Eachhydraulic motor is provided with a stationary member 90 and a rotarymember 92. As in the stationary member of the pump, and for the samereasons of improving efficiency and reducing wear in the forwarddirection of rotation, the axis of the cylinder bore is inclinedrelative to a radial line. However, inasmuch as the motor has the higherpiston forces during the inward stroke of the piston, while the pump hasits higher forces on the outward stroke, the direction of inclinationfor the motor is opposite the pump, i.e., opposite to the direction offorward wheel rotation. This is shown as angle β in FIG. 5. Rotarymember 92 rotates about wheel axis 94 in the direction of the arrow whenthe vehicle is moving forward. Rotary member 92 is provided With aplurality of lobes on the outer periphery surface which cooperate withplurality piston assemblies 98 which reciprocate in cylindrical boresformed in the stationary member.

The rotary member 92 is provided with fourteen lobes in the preferredembodiment shown, and eight piston assemblies which would cause themotor volume to be approximately double the pump volume, provided thedisplacement per cylinder is substantially the same. The motor 86 isdesigned to run in either of two modes; a synchronous mode; and a hightorque mode. During the synchronous mode half of the piston assembliesare deactivated and shifted radially outwardly so as to not engage therotary member periphery. With only four pistons actively functioning,the rotation of the hydraulic motor results in each of the four pistonsreciprocating fourteen times, generating fifty-six pressure and inletpulses. The piston assemblies and their respective bores define enclosedvolumes similar to that of the hydraulic pump which are connected totiming valves 100 by a fluid passageway not shown. In the synchronousmode, the displacement of the pump and the displacement of the motor aresubstantially the same, causing the mechanically driven wheel and thehydraulic driven wheel to turn at the same speed. Since there willinvariably be some leakage, preferably the motor displacement isslightly less than the pump displacement. Preferably, the pumpdisplacement is 2% to 8% greater than the motor displacement, and mostpreferably, the pump displacement is 5% greater than the motordisplacement. Not only does this displacement difference result incompensation for leakage but it helps balance the road load tractiveforce out among all of the vehicle wheels.

In a non-synchronous mode, all of the eight motor pistons activelyengage the rotary member effectively doubling the size of the motordisplacement. This causes the hydraulically driven wheels to turn athalf of the speed of the mechanically driven wheels. The motor will gointo the non-synchronous mode in rare occasions, such as when a vehicleis stuck in the snow and the mechanically driven wheels are unable toachieve traction. By effectively increasing the displacement of themotor, the hydraulically driven wheels can be turned at a reduced speedat much higher torque, enabling the vehicle to free itself fromconditions in which it can not do in the synchronous mode. The motor isshifted between the synchronous and non-synchronous mode by thresholdvalve 102. The non-synchronous mode can be entered into automatically asa function of line pressure or manually by controlling the magnitude ofthe lube pressure. The threshold valve 102 is conveniently locatedwithin the cylinder block adjacent the piston assembly which itregulates.

The construction of a channel plate and the cylinder block which make upthe stationary member of the rotary motor is shown in more detail inFIGS. 6-9. Channel plate 50 is generally semi-circular in shape having aL-shaped cross-section, as shown in FIG. 7. The channel plate isprovided with a planar face 104 which sealingly cooperates with thecylinder block 52 which is affixed thereto. Machined or cast into theface of the channel plate are three generally arcuate shaped groovesproviding fluid passageways. A groove 106 provides a lubrication fluidchannel. Another groove 108 provides a pressure A channel, and a thirdgroove 110 provides a pressure B channel.

When the vehicle is driving forward, the pressure in A channel is highand the pressure in B channel is low. In reverse, pressure A channel islow and the pressure B channel is high. Channels 106, 108 and 110 extendcircumaxially about the face 104 to cooperate with corresponding portsformed in the cylinder block accompanying each of the timing valves. Allthree of the fluid channels are provided with an outlet, such as Bchannel output 112, shown in FIG. 7, which couples the rear motor withthe corresponding pump. Grooves 64, which serve to align the pistons andtiming valve, are machined in the channel plate, as shown in FIG. 6.Cylinder block 52 shown in FIGS. 8 and 9 is provided with a face 114which sealingly engages face 104 of channel plate 50. Cylinder block 52is provided with cylindrical bores 56 for the piston assemblies, andbores 116 for the timing valve assembly 78. Each timing valve cavity isconnected to two cylindrical bores containing piston assemblies by wayof fluid passageways 119 and 120. These fluid passageways arealternatively coupled either to pressure B channel or pressure Achannel, dependent upon the position of the, timing valve. Ports 122 and124 extend between the timing valve bore and channels A and Brespectively.

A lubrication channel 106 in the channel plate is coupled to the variouspiston assemblies and valves which reciprocate in the cylinder block. Asshown in FIG. 9, cylinder bore 156 is coupled to lube channel 106 bylube port 126. Preferably, cylinder bore 156 is provided with a pair ofO-ring grooves 128 and 130 axially spaced on opposite sides of lube port126. The lubrication region bounded between the O-rings also serves as abuffer to limit the pressure exerted on the outermost seal, O-ring 128.

The hydraulic wheel motor and pump system is schematically shown in FIG.10. Right front hydraulic pump 142 is shown connected to the left rearhydraulic motor 88. Also depicted schematically are the A and B pressurelines 132 and 134 which hydraulically interconnect the motor and pump.FIG. 10 also schematically shows the arrangement of the lubricationsystem 136 and actuator system 138.

The actuator circuit is used to turn the four wheel drive system on andoff, either manually by driver control, or automatically by reacting tosignals from a controller which could be activated by signals such asvehicle speed, wheel slip, and engine torque. To engage the four wheeldrive, the electric motor 131 is turned on, driving charge pump 133. Thecharge pump receives fluid from oil sump 169 through passage 157, anddischarges fluid at a higher pressure through passage 135. Pressurizedfluid flows out of the pump through check valve 151 to port 155, afterwhich it is inhibited by orifice 159 so that a pressure is built up inthe port 155. The pressurized fluid acts upon valve 141, causing thevalve to overcome spring 143 and move to the left, thereby sealing offport 149, and connecting port 147 to port 145. Pressurized fluid fromthe charge pump also moves through passage 135 to chamber 163 where itcommunicates with piston 165 and also passes through restricting orifice161 and valve ports 145 and 147 to chamber 175 where it communicateswith piston 173. Piston 165 and piston 173 are connected by shaft 167and directly oppose each other. However, inasmuch as piston 173 islarger than piston -65, equal pressures upon the two pistons causes theforce of the larger piston to overcome the force of the smaller pistonand springs 171, moving the piston to the left. Pressure in line 135 isultimately controlled by a pressure relief valve made up of ball 137 andspring 139, the valve opening at a predetermined pressure anddischarging excess pump fluid capacity to the sump. Pressurized fluidfrom passage 135 now moves through passage 187 to the lube channel ofthe right wheel pump and left wheel motor. Fluid then flows through thelube circuit of each pump to each timing valve and through theselect-low valve to pressure channels A and B, causing all pump pistonsand timing valves to move inward to engage the rotating cam ring.Pressurized fluid in channels A and B also moves through longitudinalpipes to the connecting wheel motor and through orifice 191 to the motorlube circuit. This causes the motor pistons and timing valves to moveinward to engage their cam ring.

To disengage the four wheel drive system, the electric motor is turnedoff, causing the charge pump to stop rotation. Flow through orifice 159causes a decrease in pressure in line 135 and chamber 155. Spring 143causes valve 141 to move to the right, closing port 145 and connectingport 147 to port 149. As soon as port 147 is connected through port 149and channel 157 to sump 169, pressure is quickly reduced in chamber 175which corresponds with piston 173. Spring 171 moves pistons 173 and 165to the right. As piston 165 strokes to the right, pressure in chamber163 decreases to atmospheric, causing check valve 151 to close. Spring171 continues to exert a force to move the pistons 163 and 171 to theright, decreasing pressure in chamber 163, and lines 135, 187, and 189to below atmospheric. This negative gage pressure in the pump lubecircuit also lowers the pressure in channels A and B through theselect-low valve. In the motor, with pressure in lines A and B belowatmospheric, check valves 193 and 195 open to exhaust motor lubepressure. With all pressures in both the wheel pump and wheel motorbelow atmospheric, external atmospheric pressure moves all pistons andtiming valves outward, disengaging them from the cam rotors and allowingthe rotors to spin freely.

To prevent any disproportionately high pressures from arising betweenthe left and right motor circuits in sharp cornering maneuvers, apassage 183 with restricting orifice 185 connects to chamber 177 in thetwo motor housings. Double acting check ball 181 insures that chamber177 connects to the higher pressure of channel A and B.

Four threshold valves 102 are used to control pressure to the fourauxiliary pistons in each motor. In its normal off position shown inFIG. 23, spring 103 holds the valve in its inward position, connectingport 113 to piston feed passage 117. The outside of the valve has threediametral sizes: largest diameter 105; middle sized diameter 107; andsmallest diameter 111. Port 109 connects with the output of itscorresponding timing valve, generating a force unbalance from thedifference in areas of diameter 105 and diameter 107. Adding to thisforce unbalance is lube pressure entering from channel 106 through port113 to chamber 115 and acting on diameter 107. The combined forceunbalance thus generated is opposed by spring 103, and when the force issufficiently large, the valve moves outward against the spring 103. Whenthe valve moves outward, port 109 now corresponds with area differenceof diameter 105 and diameter 111, causing a snap action in the valvemotion. Passage 101 from the timing valve now connects with pistonpassage 117, causing the motor displacement to increase correspondingly.For the four threshold valves in each motor, the spring forces can allbe equal to make the four valves act in concert, or the springs forcescan be sequentially incremented to bring the auxiliary motor pistons inin a stepped sequence.

A detailed drawing of the piston assembly is shown in FIGS. 11 and 12.Piston 54 has a generally cylindrical surface 140 for cooperating withthe cylindrical bore of the cylinder block. Attached to the lowerportion of the piston assembly is fork 142 which supports shaft 146 androller follower 60. Roller follower 60 is mounted on shaft 146 supportedby needle bearinq assemblies 148. Alignment roller 62 is pivotablysupported on stub shaft 150 to maintain the alignment of the rollerfollower with the cam as previously described.

A piston assembly 54 is provided with a lubrication system and aninternal pressure relief valve. The outer cylindrical surface 140 isprovided with a lubrication groove 152, which during operation isgenerally aligned between O-ring grooves 128 and 130 in the cylindricalbore 156. An internal tubular passageway 154 connects groove 152 tobearings 148 as shown.

Piston assembly 54 has a raised head 156, as shown in FIGS. 11 and 12.Piston head 156 is provided with a pressure relief port 158. Thepressure relief port is connected to groove 152 and the lubricationsystem. The pressure relief port is maintained in a normally closedposition, as shown in FIG. 12 by ball 160 and spring 162. In the eventan over pressure condition occurs which could potentially result indamage to the system, ball valve 160 will move off its seat to the openposition, as shown in FIG. 11, allowing hydraulic fluid to flow from thechamber through the pressure relief port 158, to be exhausted into thelubrication system.

A preferred embodiment of the timing valve assembly 78 is shown in FIGS.13 and 14. The timing valve assembly is made up of three basic parts: aspool valve 164; a body 166; and a follower 168. Body 166 is fabricatedof a number of subparts which are pressed together for the ease inmanufacturing, and define inlet/outlet port 170 and 170' which arecoupled to two piston cavities by way of fluid passages in the valvebody, such as passageways 118 and 120, shown in FIG. 8. Depending uponthe position of the valve spool relative to the body, inlet/outlet port170 is alternatively coupled to A port 172 or B port 174 whichcommunicate with A port 122 and B port 124 in cylinder block 52 whichare, in turn, coupled to A channel and B channel 108 and 110 in channelplate 150. The valve spool 164 has a central narrow region and tworelatively larger diameter cylindrical regions on opposite ends thereto.The narrow region is of an axial length which only permits inlet/outletport 170 to be coupled to one of ports 172 or 174. In the raisedposition, shown in FIG. 13, the A port 172 is blocked and the B port 174is open. Preferably, elastomeric seals 176 are provided to preventleakage between the valve spool and body, as shown. Elastomeric seals176 are axially retained by the body structure, and are radiallyretained and positioned by O-ring 178 and 178', The O-ring helpsmaintain proper radial alignment of elastomeric seal relative to thespool. The upper portion of the valve body, in conjunction with thevalve spool end, define an enclosed cavity 180 which is coupled to thelubrication system. The lubrication system pressure tends to bias thevalve spool to the extended position, thereby maintaining the followerin contact with rotary member 46. Preferably, the valve spool isprovided with a vent port 182 which extends from the cavity 180 to theregion of the spool periphery below A port 172. In the event there isany leakage of high pressure fluid past the valve body seal adjacent theA port, vent 182 will allow the fluid to return to the lubricationsystem, thereby minimizing leakage in the maximum pressure load, whichwill be seen by spool 194. Spool seal 194 is a conventional design,since it continually maintains contact with body and spool.

Valve spool follower 168 as shown in the preferred embodiment, is formedof a plastic material and pressed onto the valve spool. Since the valvespool follower loads are relatively low compared to the pistons, aroller follower is not necessary. Preferably, the radius of the followersurface in contact with the rotary member has a substantially similarradius as the roller follower of the piston assembly, so that the valvelift versus time curve is substantially the same shape as that of thepiston assembly. Follower 168 is also provided with an alignment pin 186which cooperates with a groove 64 in the channel plate to prevent valvespool and follower rotation.

Piston and valve spool timing curves are shown in FIG. 15, which is aplot of lift versus time. Additionally, piston velocity is shown, sincepiston velocity is directly related to the flow rate of hydraulic fluidexiting or entering the chamber. Piston lift velocity are depicted bycurves 188 and 190. Valve position is depicted by curve 192. The valveand piston lift curve are intentionally out of phase by an amountapproximately equal to about one-half of the lift duration. As can beclearly seen from the curve shapes in FIG. 15, the piston and timingvalve motion is not sinusoidal, rather there is a dwell period at boththe crest and root of the curve and a constant velocity segment half waybetween. Piston velocity curve 190 is, in essence, a first derivative ofpiston position curve 188. During the dwell periods when piston is atthe bottom and peak of its travel, piston velocity is zero, as shown inthe curve segment indicated by double-ended arrow 194. It is during thedwell periods when piston velocity is zero and fluid flow is likewisezero that the valve spool shifts, causing the inlet/outlet port 70 to becoupled to the opposite port. For example, in a pump, when the piston isrising, the fluid inlet/outlet port is connected to the A port, and whenthe piston is descending the inlet/outlet is connected to the B port. Inthe motor, the motion would be just the opposite; when a piston isdescending the inlet/outlet would be coupled to the A port, and when thepiston is rising, the inlet/outlet would be coupled to the B port.

The flow output of the pump in A-line 134 is a composite of the pressurepulses from all of the cylinders. Similarly, the input to the pump fromB-line 132 is a composite of all the input strokes from the variouscylinders. The cam profile and follower construction, is preferablydesigned so that all of the individual output pulses cumulatively form asubstantially uniform flow rate. As can be seen in FIG. 16, the flowoutput for b 1/7th of a revolution yields four discrete pulses, 196,196', 196" and 196'", which can be cumulatively added to provide asteady flow rate 198. There are only four pulses in a 1/7th of arevolution, as opposed to eight, since each pulse represents the outputof two cylinders similarly timed. The flow input and flow output of boththe motor and the pump are both substantially uniform, therebyeliminating the cost and energy losses associated with an accumulator.

As previously described, the preferred motor has dual displaced modes; asynchronous mode in which the motor and pump rotate at substantiallysame speeds, and a non-synchronous mode which the pump turns faster thanthe motor. As can be seen in FIG. 16, the torque of the hydraulicallydriven wheel pair is directly related to pressure, as one would expect.This relationship is substantially linear. The higher the pressure, thehigher the wheel torque. In a preferred embodiment, threshold valve 102is pressure activated automatically when the line pressure exceeds 2500psi. When threshold valve 102 opens, the motor displacement increases byan amount proportional to the displacements of the added cylinders,resulting in the step in a tractive force curve 200. As previouslyindicated, the threshold valve could alternatively be manually operatedby the vehicle occupant using the alternate manual driver input, whichis shown schematically in FIG. 10.

The beneficial effect of the hydraulic pump and motor four wheel drivesystem on a vehicle's performance is schematically shown in FIG. 18.Curve 202 is a plot of total tractive force versus friction coefficientfor a two wheel drive vehicle having either one of its primary drivewheels on a low friction surface. In the four wheel drive mode, curve204 represents the total tractive force available when one of theprimary drive wheels is on a low coefficient surface and the remainingthree wheels are on a high coefficient surface. When one entire side ofthe vehicle is on a low coefficient surface in the four wheel drivemode, as shown in curve 206, tractive force is slightly lower than curve204, but substantially higher than the two wheel drive condition. Whenall four wheels are on a low coefficient friction as shown in curve 208,tractive force is still substantially better with the hydraulic fourwheel drive system employed.

The hydraulic four wheel drive device works when a vehicle is operatingin reverse as well as in forward, as previously mentioned. When thevehicle is running in the forward mode, A pressure is high and Bpressure is low. In reverse, A pressure becomes low and B pressurebecomes high. Since the lubrication system is directly coupled to the Bpressure in the forward mode, when the vehicle shifts into reverse, itis necessary to couple the lubrication to the A pressure line to avoidfull line pressure from entering into the lubrication system. A lowselect valve 210 connects the lubrication line 136 alternatively to Apressure line 134 or B pressure line 132, as shown in FIG. 10.

An enlarged, low select valve is shown cut-away in FIG. 19. The valveconsists of a body 12 and a spool 214. The valve body has two inlets, Ainlet 216, and a B inlet 218. The body is also provided with twooutlets, an A outlet 220 and a B outlet 222. The spool shifts between anA position shown, in which the A inlet is coupled to the A outlet whenpressure B is higher than pressure A, to a B position, where the valvewould be moved to the extreme right and the B inlet would be coupled tothe B outlet, and the A outlet would be closed. The spool automaticallyshifts to the appropriate position so that the outlet to the lubecircuit will be coupled to the fluid inlet having the lowest pressure.

FOUR WHEEL TORQUE CONTROL

The concept of independent torque control for a four wheel drive vehicleis currently the area of interest in the automotive engineeringcommunity. In order to achieve maximum directional stability, a

vehicle maneuver is desirable to effectively, utilize all of the loadbearing capacity of each of the vehicle's wheels. Typically, in any typeof turning maneuver, the vehicle will lose directional stability whenone of the vehicle's four wheels exceeds the allowable wheel load,resulting in uncontrolled slippage. The concept of independentlycontrolling wheel torques is discussed in The Society of AutomotiveEngineer's Technical Paper 880701, Donald Margolis et al. DirectionalStability Augmentation for All Wheel Drive Vehicles, February, 1988,which is incorporated by reference herein.

FIG. 20 shows a plot of wheel load vectors for each of the four wheelsof the vehicle, where cornering force is the Y-axis, and lateral forceis the X-axis. The four vectors designated RF, IF, RR and LR representthe load on the four vehicle tires; right front, left front, right rearand left rear. The tire roadway breakaway load curves are depicted bysemicircular lines 224-226 which show the breakaway load for each of thevehicle tires. Note that the breakaway load for each wheel is differentas a result of varying normal forces due to weight distribution andlateral acceleration. The four vectors shown depict the maximumcornering force situation without torque control. Any additional lateralacceleration will cause the load vector RR to exceed the limits of curve228 resulting in the right rear wheel breaking loose and the entirevehicle to spin out of control. Additional lateral acceleration could beachieved, however, if the other three wheels could be more effectivelyutilized. Double ended arrows 232, 234 and 236 represent the unusedcapacity of the right front, left front and left rear tires,respectively.

An alternative embodiment of the present invention is shown in FIGS. 21and 22 which utilizes a low pressure bias pump 238 and a high pressurebias pump 240 to transfer fluid pressure from one side of the vehicle tothe other. For example, when the vehicle shown in FIG. 21, makes a leftturn caused by the rotation of the front wheels through an angle α, α'the four tires will be loaded in a manner similar to that shown in FIG.20. In order to maximize lateral acceleration, the high pressure biaspump 240 will be driven by motor 242 in the direction which would causea transfer of fluid from the left rear hydraulic motor to the right rearhydraulic motor.

The hydraulic pump is coupled to the line connecting the right and leftrear wheels on opposite sides of a restriction 244 which enablessubstantially different pressures to be applied to each of the hydraulicmotors. Since the hydraulic fluid is substantially incompressible, verylittle actual fluid needs to be transferred from one motor to the otherto cause a substantial change in pressure and resulting tractive force.The high pressure pump motor 242 and corresponding low pressure pumpmotor 246 can be rotated in either direction in response to a controlsignal from a microprocessor 248. Microprocessor 248 regulates theoperation of both bias pumps dependent upon driving conditions andreceives inputs from steering angle, lateral acceleration vehiclevelocity time and drive torque from sensors 250, 251, 252, 253 and 254of a conventional design. The bias pumps transfer torque side to side,however, there is some front to rear effect as a result connection ofthe pumps and motors.

To further control front to rear distribution if necessary bypass valves256 and 258 may be installed in the hydraulic lines connecting the motorand pump to allow some of the fluid to short circuit or bypass the pumpwhich effectively transfer load to the front mechanically driven wheel.Once again, like the bias pump, very little fluid needs to betransferred to affect a significant torque change. Utilize front andrear bias pumps in combination with bypass valves, the torque of eachwheel can be independently controlled during a vehicle maneuver tooptimize tractive force and directional stability. The hydraulic drivesystem readily lends itself to independent control.

It is also understood, of course, that while the form of the inventionherein shown and described constitutes a preferred embodiment of theinvention, it is not intended to illustrate all possible forms thereof.It will also be understood that the words used are words of descriptionrather than limitation, and that various changes may be made withoutdeparting from the spirit and scope of the invention disclosed.

I claim:
 1. A hydraulic drive apparatus for a motor vehicle having apair of left and right mechanically driven wheels, and a pair of leftand right non-mechanical driven wheels, said hydraulic drive apparatuscomprising:a right and left rotary hydraulic pump coupled to the rightand left mechanically driven wheel respectively, each of said pumpshaving an A and B port providing a fluid output and a fluid inputrespectively when the vehicle is driven in the forward direction, and afluid input and a fluid output respectively when the vehicle is drivenin reverse; a right and left rotary hydraulic motor coupled to the rightand left non-mechanically driven wheel respectively, each of said motorhaving an A and B port providing a fluid input and fluid outputrespectively when the vehicle is driven in the forward direction, and afluid output and fluid input respectively when the vehicle is driven inreverse; conduit means for hydraulically coupling the A and B ports ofthe right rotary hydraulic pump to the A and B ports of the left rotaryhydraulic motor respectively, and hydraulically coupling the A and Bports of the left rotary hydraulic pump to the A and B ports of theright rotary hydraulic motor respectively, forming a pair ofsubstantially independent closed-loop hydraulic circuits; therebymaximizing the vehicle's tractive force when one side of the vehicle isplaced on a low friction coefficient surface; and means for by-passingsufficient hydraulic fluid between the two closed-loop hydrauliccircuits to enable the vehicle to make tight turns without wheelslippage.
 2. The invention of claim 1 wherein said means for by-passingfurther comprises a first and a second fluid passageway connecting theclosed-loop hydraulic circuit A ports together and the B ports togetherto allow limited fluid flow therebetween.
 3. The invention of claim 2further comprising a pair of high pressure select valves cooperatingwith the A and B ports of each of the closed-loop hydraulic circuits andthe first to selectively couple the pump inputs together, and a pair oflow pressure select valves cooperating with A and B ports of each of theclosed-loop hydraulic circuits and the second fluid passageway toselectively couple the pump inputs together, said first and said secondfluid passageways providing a high pressure and a low pressure fluidpassageway respectively.
 4. The invention of claim 3, wherein said highpressure fluid passageway is restricted limiting the flow of fluidtherethrough, while said low pressure passageway is relativelyunrestricted.
 5. The invention of claim 4, wherein said high pressureselect valves further comprises a double acting check flow valve havinga first inlet coupled to an A port, a second inlet coupled to a B port,and an outlet coupled to the restricted high pressure fluid passageway,said check flow valve automatically coupling the inlet having the higherpressure to the outlet, so that the restricted fluid passageway connectsthe pump outputs in both the forward and reverse drive modes.
 6. Ahydraulic drive apparatus for a motor vehicle having a pair of left andright mechanically driven wheels, and a pair of left and rightnon-mechanical driven wheels, said hydraulic drive apparatuscomprising:a right and left rotary hydraulic pump coupled to the rightand left mechanically driven wheel respectively, each of said pumpshaving an A and B port providing a fluid output and a fluid inputrespectively when the vehicle is driven in the forward direction, and afluid input and a fluid output respectively when the vehicle is drivenin reverse; a right and left rotary hydraulic motor coupled to the rightand left non-mechanically driven wheel respectively, each of said motorshaving an A and B port providing a fluid input and fluid outputrespectively when the vehicle is driven in the forward direction, and afluid output and fluid input respectively when the vehicle is driven inreverse; a first pair of A and B conduits hydraulically coupling the Aand B ports of the right rotary hydraulic pump to the A and B ports ofthe left rotary motor respectively, and a second pair of A and Bconduits hydraulically coupling the A and B ports of the left hydraulicrotary pump to the A and B ports of the right rotary hydraulic motorrespectively, forming the substantially independent closed-loophydraulic circuits; a low pressure fluid passageway and a restrictedhigh pressure fluid passageway providing limited fluid flow between thetwo closed-loop hydraulic circuits; and a pair of high pressure selectvalves cooperating with the A and B ports of each of the closed-loophydraulic circuits and the high pressure fluid passageway to selectivelycouple the pump outputs together, and a pair of low pressure selectvalves cooperating with the A and B ports of each of the closed-loophydraulic circuits and the low pressure fluid passageway selectivelycoupling the ports acting as pump inputs together, thereby allowinglimited fluid flow between the two closed-loop hydraulic circuitssufficient to enable the vehicle to make tight turns without wheelslippage yet providing tractive force when one side of the vehicle isplaced on a low friction coefficient surface.